Axial Piston Compressor

ABSTRACT

Axial piston compressor, especially a compressor for the air-conditioning system of a motor vehicle, having a housing and, for drawing in and compressing a coolant, a compressor unit arranged in the housing and driven by means of a drive shaft, the compressor unit comprising pistons, which move axially back and forth in a cylinder block, and a tilt plate (swash plate, wobble plate or tilt ring) which drives the pistons and rotates together with the drive shaft, wherein the tilt plate is so constructed or mounted that its tilting behaviour acts in a self-limiting manner such that at high speeds of rotation of the compressor, especially at very high speeds of rotation or at the maximum speed of rotation, the angle of maximum deflection of the tilt plate is less than the angle of maximum deflection α max  at low speeds of rotation of the compressor.

The invention relates to an axial piston compressor, especially to a compressor for the air-conditioning system of a motor vehicle, in accordance with the preamble of claim 1.

DE 195 14 748 C2 explains the tilting moments which act on a tilt plate in the case of, in principle, all compressors existing in the prior art, and also those compressors that have entered mass production, and which decisively contribute to the tilting behaviour of the tilt plate. The influences acting as moments about the centre of tilt of a tilt plate are, in detail, the following moments, the direction of the particular moments being given in brackets, with (−) denoting down-regulation (in the direction of minimum stroke) and (+) denoting up-regulation (in the direction of maximum stroke):

-   -   moment due to gas forces in the cylinder spaces (+)     -   moment due to gas forces from the drive mechanism chamber (−)     -   moment due to a restoring spring (−) (in the direction of         minimum stroke)     -   moment due to an advancing spring (+) (in the direction of         maximum stroke)     -   moment due to rotating masses (−); including moment due to         location of centre of gravity (for example, tilt plate: tilt         location≠mass centre of gravity): can be (+) or (−)     -   moment due to masses moved in translation (+)

The moment due to the rotating masses (hereinafter referred to as M_(SW)) generally has a down-regulating action; it is only in the region of very small tilt angles that as a result of, for example, an outwardly displaced location of centre of gravity (Steiner component when calculating the moment of deviation J_(yz)) an up-regulating moment can be produced in the case of the tilt plate (see DE 195 14 748 C2 in this regard). The component of the moment of deviation J_(yz) without a Steiner component already predominates, then, in the region of small tilt angles, and as the tilt angle increases the tilt plate has an increasingly down-regulating tilt moment.

Referring to DE 195 14 748 C2, a curve for the masses moved in translation having, as already explained, an up-regulating action is also shown.

Also of interest is the sum of the moments, which is likewise represented in graph form (FIG. 4 of DE 195 14 748 C2). For the entire tilt angle range, the drive mechanism exhibits up-regulating behaviour because the masses moved in translation dominate the regulation behaviour in all ranges.

From EP 0 809 027 A1 there is known a further drive mechanism which is characterised in that the delivery quantity of the compressor is so compensated by the dynamic behaviour of the drive mechanism of the compressor that the delivery quantity can be kept constant. Specifically, the following is stated therein in this respect: “In order to keep constant the delivery quantity at changing rotation speeds, use can be made of the restoring torque of the wobble plate counteracting its inclined position due to dynamic forces on the co-rotating plate part.”

Building on EP 0 809 027 A1, measures are known from DE 198 39 914 A1 as to how regulation behaviour of such a kind (at least partial compensation of the quantity delivered) can be achieved. It is proposed that the component mass of the tilt plate be so dimensioned in relation to the masses moved in translation that the centrifugal forces of the tilt plate influence its regulation behaviour. Specifically, it says in that respect that the rotating mass of the drive plate is greater than the total mass of all the pistons so that the centrifugal forces occurring during rotation of the drive plate are sufficient to deliberately counteract the pivoting movement of the drive plate in regulating manner and thereby to influence, especially to reduce or to limit, the piston stroke and, accordingly, the quantity delivered.

In the as yet unpublished DE 103 24 393 belonging to the Applicant it is explained why the component mass should not be the preferred parameter for influencing as desired the regulation behaviour of the drive mechanism as a result of changes in the speed of rotation.

The desired regulation behaviour of the compressor is primarily achieved not by means of the component mass of the tilt plate in relation to the masses moved in translation but rather by taking into account the mass moment of inertia of the component unit that is the “tilt plate”, which depends more on its geometry than on the component mass. A central idea is accordingly that, in the case of fluctuations or changes in the speed of rotation, the moment due to masses moved in translation is directly compensated, or even over-compensated, by the moment due to rotating masses.

In the case of such compressors it is desirable to reduce the frequency and intensity of regulatory interventions to a low level. With reference to the prior art and DE 195 14 748 C2 it is extremely disadvantageous that, when there is an increase in the delivery volume due to an increase in the speed of rotation, an additional increase in the delivery volume due to an increase in the tilt angle of the tilt plate is added thereto.

This effect has to be compensated by appropriate regulatory interventions. This is not convenient, is onerous and reduces the efficiency (fuel consumption).

EP 0 809 027 A1 specifies the goal that is to be achieved: keeping constant the delivery quantity.

It can, however, be demonstrated very simply that this is not possible solely by means of the advancing moment (down-regulation) acting on the tilt plate.

The delivery volume is directly proportional to the speed of rotation, that is to say if the speed of rotation doubles, the delivery volume also doubles.

On the other hand, the following equation applies for the tilt plate tilt moment, which is produced by the relevant moment of deviation:

M_(SW)=J_(yz)ω²

Because it is the square of the speed of rotation that influences the tilt moment, the expressed aim of “keeping constant the delivery quantity” cannot be achieved solely as a result of the design or dimensioning of the tilt plate.

The problem of the invention is to provide a compressor having optimised regulation behaviour wherein, at high and highest compressor speeds of rotation, a full-load condition is avoided.

The problem is solved by a compressor having the features in accordance with patent claim 1, preferred developments and embodiments being described in the subordinate claims.

It is accordingly a fundamental point of the invention that the tilting behaviour of the tilt plate acts in a self-limiting manner such that at high speeds of rotation of the compressor, especially at very high speeds of rotation or at the maximum speed of rotation of the compressor, the angle of maximum deflection of the tilt plate is less than the angle of maximum deflection α_(max) at low speeds of rotation of the compressor. A measure of such a kind rules out the possibility of the compressor's operating with too great a delivery volume at high speeds of rotation, especially at its maximum speed of rotation. Because this measure intervenes in the delivery volume in regulating manner, the number of regulatory interventions required is reduced (tilt plate compressors are usually regulated by means of the pressure difference between the pressure prevailing in the drive mechanism chamber of the compressor and the pressure prevailing on the output side of the compressor, that pressure difference governing the piston stroke and, therefore, the delivery volume of the compressor). This also results in the fact that in particular at moments of increasing speed of rotation, especially in the region of high speeds of rotation, a markedly reduced number of power-consuming regulatory interventions or possibly no regulatory intervention at all is/are required, which particularly in the case of use of a compressor in motor vehicles allows the power at those moments to be directed towards driving the motor vehicle without reduction. This is very advantageous especially in the case of overtaking manoeuvres.

Preferably, a compressor according to the invention is so constructed that both the geometry and also the dimensioning of all parts of the compressor that are moved in translation and of all parts moved in rotation are such that for predetermined tilt angles of the tilt plate, especially between a predetermined minimum tilt angle and a predetermined maximum tilt angle, the moment M_(k, ges) due to the masses moved in translation is so selected as to be less than the moment M_(SW) due to the moment of deviation, that is to say is so selected as to be less than the moment due to the mass inertia of the tilt plate, that at high speeds of rotation of the compressor, especially at very high speeds of rotation or at a maximum speed of rotation, the angle of maximum deflection of the tilt plate is less than the angle of maximum deflection at lower speeds of rotation of the compressor. Those parts of the compressor which are moved in translation include the pistons and also, where appropriate, the piston rods and sliding blocks or the like. The parts moved in rotation include, basically, the tilt plate and also, where appropriate, one or more members for transmitting its drive. This arrangement constitutes an embodiment of a compressor according to the invention which is simple to produce in structural terms and accordingly economical to manufacture. In addition, of course, the advantages already mentioned hereinbefore are also entirely applicable to the present preferred embodiment, especially the advantage of a low number of regulatory interventions necessary during operation of a compressor according to the invention.

It should be mentioned at this juncture that modern axial piston compressors, especially compressors intended for use in a motor vehicle air-conditioning system, bring about, under partial loading, a coolant mass flow—corresponding as it were to the compressor output—of a magnitude such that a sufficiently high cooling performance is made available. Particularly at high, very high and also at the highest speed(s) of rotation, the compressor therefore has to be down-regulated, the number of regulatory interventions required being very greatly reduced, or no regulatory intervention at all being required, as a result of the present invention. Reference should also be made, moreover, to the safety aspect, which plays an important role especially in the case of compressors which are operated with a high-pressure coolant. The arrangement according to the invention prevents a high speed of rotation and a large compressor piston stroke from occurring at the same time, which may well, in the case of compressors according to the prior art, result—given the (possibly sudden) occurrence of a high speed of rotation and a large stroke—in an increase in compressor pressure such that bursting of the compressor housing ensues.

The compressor according to the invention can comprise at least one arrangement which exerts an adjusting force on the tilt plate. Preferably, said arrangement is associated with the compressor in addition to an arrangement for adjustment or control or regulation of its drive mechanism chamber pressure. This arrangement or these arrangements can function on its/their own as the self-limiting element of the tilt plate but can also be arranged in a compressor wherein the distribution of masses or the distribution of moments of the tilt plate already brings about down-regulation thereof at high speeds of rotation. In this case, an already existing tendency of the compressor, due to the geometry of the tilt plate mechanism, to down-regulate at a high speed of rotation is assisted or reinforced. The at least one arrangement for exerting an adjusting force optionally comprises a resilient element, which can be present especially in the form of a restoring spring, and/or an actuator or an adjusting piston. Alternatively or additionally, a compressor according to the invention can comprise an adjusting member which is dependent on centrifugal force and/or a further spring. It is also feasible for the at least one arrangement for exerting the adjusting force to comprise a throttling location, the cross-section of which can be varied by means of an adjusting member, especially by means of an adjusting piston. The throttling location can be arranged especially on the intake side or on the pressure side of the compressor. If the throttling location is arranged on the intake side of the compressor, it is accordingly possible, as a result, for the density of the intake gas to be regulated or reduced as desired. As an adjustment variable for the adjusting member, especially an adjusting piston, preference is given to the pressure prevailing in the drive mechanism chamber of the compressor and/or the pressure prevailing on the outlet side of the compressor. These adjustment variables are readily accessible. In a further preferred embodiment, the adjusting member is regulated or controlled internally, especially by means of a solenoid or like apparatus. Alternatively, a constant cross-section of the throttling location is also feasible. All the embodiments mentioned hereinbefore are structurally simple measures, which are accordingly economical to produce. Depending on the area of use of the compressor and the desired adjusting force or restoring force, the various measures can be implemented either singly or also in combination with one another in the compressor, as a result of which it is possible to achieve a desired regulation behaviour by simple means.

In a preferred embodiment, the centre of gravity of the tilt plate is located on the tilt axis thereof. The tilt axis of the tilt plate is in turn preferably located on the central axis of the drive shaft. These structural measures ensure quiet running of the compressor with a low degree of wear and with minimal imbalance.

In a further preferred embodiment, the tilt plate is ring-shaped, that is to say is in the form of a tilt ring. This ensures an optimum ratio between the mass of the tilt plate or tilt ring (which is low) and its mass moment of inertia (which is high).

Preferably, the angle of maximum deflection of the tilt plate in the region of low and medium speeds of rotation of the compressor corresponds to an angle α_(max), whereas the angle of maximum deflection at the maximum speed of rotation of the compressor corresponds approximately to an angle of α_(max)/2. Under the terms of the present Application, regions of low and medium speeds of rotation extend up to about half the maximum speed of rotation of the compressor. In the case of compressors of modern design, which reach a maximum speed of rotation of about 8000 to 10,000 revolutions per minute, a range of low and medium speeds of rotation accordingly extends up to about 4000 revolutions per minute, although it should be pointed out that the above-mentioned figures are of course to be regarded as merely being examples.

Preferably, for speeds of rotation from about 4000 to 5000 revolutions per minute upwards, the tilt plate of a compressor according to the invention should no longer be capable of being pivoted up to an/the angle α_(max). This ensures optimum loading limitation, which has advantageous effects particularly from the safety aspect. Both the danger of icing-up and also the danger of the compressor housing bursting are markedly reduced compared to compressors according to the prior art.

In a further preferred embodiment, the limitation of the pivot angle becomes greater for increasing speeds of rotation of the compressor. For this embodiment, too, there are obtained advantages, especially when seen from the safety aspect, which are analogous to the above-mentioned embodiments.

The ratio of the moment of deviation in the y direction to the total mass of all parts moved in translation J_(y)/m_(k, ges), which include as explained hereinbefore, for example, the pistons and also, where appropriate, the sliding blocks and piston rods, is preferably at least about 1000 mm². Even more advantageous is a ratio J_(y)/m_(k, ges) of more than 1500 mm². The ratio of the moment of deviation in the z direction to the total mass of all parts moved in translation J_(z)/m_(k, ges) is, in a further preferred embodiment, at least about 2000 mm²; a value of more than 3000 mm² is even more advantageous. As a result of the large ratio between the moment of deviation and the total mass of the parts moved in translation, the desired regulation characteristic curve is achieved in simple manner.

In accordance with the invention it has also been recognised that a “fitting” spring constant should be selected (this governs the slope of the regulation characteristic curve). In a further preferred embodiment, a restoring spring which exerts a restoring force on the deflected tilt plate has a spring constant of less than about 60 N/mm. Even more advantageous is using restoring springs which have a spring constant of less than 30 N/mm, the moments of the parts moved in translation and in rotation of course being appropriately matched in each particular case. Springs having such spring constants are obtainable at low cost, require little space for installation and, in addition, are subject to a low degree of wear.

In the case of the embodiments described herein, the tilt plate has a directly defined or indirectly defined (by means of the sliding sleeve) minimum and maximum end-stop. The maximum end-stop is not reached in the case of high speeds of rotation and is prespecified in fixed, non-regulatable manner in the embodiments described hereinbefore.

In a further preferred embodiment, a compressor according to the invention has an end-stop which is dependent on the speed of rotation and which limits the angle of maximum deflection of the tilt plate in dependence on the speed of rotation of the compressor. This end-stop which is dependent on the speed of rotation serves as a further measure for ensuring that at high speeds of rotation the angle of maximum deflection of the tilt plate is less than in the case of low speeds of rotation. In this case, for example, a shoulder on the drive shaft is feasible, which shoulder prevents pivoting beyond a predetermined angle. A different arrangement, for example comprising a sleeve displaceably mounted on the drive shaft, or the like, would also be feasible. The end-stop can be controlled, for example, by means of the centrifugal force (which is dependent on the speed of rotation).

It should be pointed out at this juncture that it has been recognised in the context of the invention that the avoidance of outwardly displaced locations of the centre of gravity avoids relatively marked bends in the regulation characteristic curve.

The invention will be described hereinbelow with reference to further advantages and features by way of example and with reference to the accompanying drawings. The drawings show in

FIG. 1 a diagram of a tilt ring of a compressor according to the invention,

FIG. 2 a diagram of a tilt plate for the purpose of explaining the co-ordinate system used in the present Application,

FIGS. 3-7 diagrams showing the regulation characteristic curves of two embodiments of a compressor according to the invention, in dependence on various parameters such as speed of rotation, suction pressure, compression pressure and spring constant of a restoring spring, and

FIG. 8 a diagram showing the regulation characteristic curves of a compressor according to the invention in dependence on further (additional) parameters.

For better clarification of the invention, the tilt moment of the pivotable part of the tilt plate arrangement can be expressed in simplified form by means of equations which are relevant to a ring-shaped tilt plate component.

The simplified derivation shown hereinbelow is therefore to be regarded as being given merely by way of example; in the case of a more complex geometry of the tilt plate, the mass moments of inertia and moments of deviation and other variables influenced by the geometry and density would be calculated by means of CAD.

For the derivation of the moment of deviation the following mathematical relationships generally apply (co-ordinate system FIG. 1):

-   -   Direction angles of the x axis relative to the main inertia axes         ξ, η, ζ     -   Direction angles of the y axis relative to the main inertia axes         ξ, η, ζ     -   Direction angles of the z axis relative to the main inertia axes         ξ, η, ζ

The co-ordinate system used herein can be seen from the diagram in FIG. 2. The following also applies to a “ring”:

-   -   and

(Note: J₃≈2J₂)

Aim: J_(yz) should have a particular magnitude; necessarily increases.

For the moment of deviation, which governs the pivoting movement, the following applies:

Independently of FIG. 2, the following holds true for the moment due to mass forces of the pistons:

and also the moment M_(SW) due to the moment of deviation:

In the context of the invention, the following moment ratio should be established by structural means:

M_(SW)≧M_(k, ges) or

As already mentioned, the (tilting) moment of the tilt ring due to the associated moment of deviation can be deliberately adjusted by means of various parameters (geometry, density distribution, mass, mass centre of gravity) so that

M_(SW)≧M_(k, ges)

holds true. However, M_(SW) should especially be markedly greater than M_(k, ges).

In the context of the equations given, the variables denote the following:

θ rotation angle of the shaft (the considerations above and below being made on the basis of θ=0 for the sake of simplicity)

η number of pistons

R distance from piston axis to shaft axis

ω speed of rotation of shaft

α tilt angle of tilt ring/tilt plate

m_(k) mass of a piston including sliding block (pair)

m_(k,ges) mass of all pistons including sliding blocks

m_(SW) mass of tilt ring

r_(a) external radius of tilt ring

r_(i) internal radius of tilt ring

h height of tilt ring

g density of tilt ring

V volume of tilt ring

β_(i) angle position of piston i

z_(i) acceleration of piston i

F_(mi) mass force of piston i (including sliding blocks)

M(F_(mi)) moment due to mass force of piston i

M_(k,ges) moment due to mass force of all pistons

M_(SW) moment (or advancing moment of tilt ring/tilt plate) due to the moment of deviation (J_(YZ))

When selecting the parameters for achieving the desired moment ratio, the location of the centre of gravity should, in accordance with the invention, largely have no influence. This means that when calculating the governing moment of deviation, the Steiner component to be taken into account for the mass centre of gravity is basically

y_(s)z_(s)m=0

In this case the co-ordinate z_(s) is located in the direction of the shaft axis and the co-ordinate y_(s) is located perpendicular thereto and perpendicular to the axis of tilt x_(s) of the tilt plate (x_(s) is of course also=0).

The moment of deviation J′_(yz) including Steiner component results in

J′ _(yz) =J _(yz) +y _(s) z _(s) m

In accordance with the derivation given for a tilt plate in ring form, the variables governing the regulation behaviour were calculated. Tables 1 and 2 give the values for two different preferred embodiments of compressors according to the invention.

TABLE 1 Tilt moment determination for the tilt plate of a first preferred embodiment of a compressor according to the invention Tilt plate tilt moment determination Tilt plate variant 1

TABLE 2 Tilt moment determination for the tilt plate of a second preferred embodiment of a compressor according to the invention Tilt plate tilt moment determination Tilt plate variant 2

A sensible design criterion is, as already mentioned, the ratio J_(y)/m_(k, ges) or J_(z)/m_(k, ges), which characterises the moment ratio

M_(SW)≧M_(k,ges)

and therefore the regulation behaviour of the compressor drive mechanism independently of the coolant (which can be, for example, CO₂, R134a, R152a etc.). As can be seen from the two Tables, the preferred embodiments of a compressor according to the invention have respective ratios J_(y)/m_(k, ges) of more than 1000 mm² and more than 1500 mm², whereas the ratio J_(z)/m_(k, ges) is in both cases clearly above 2000 mm², and even about 3000 mm² in the case of the second preferred embodiment. FIG. 3 shows the regulation characteristic curves of a compressor according to the invention for the coolant CO₂ and a spring constant of the restoring spring of 30 N/mm for both a maximum speed of rotation (8000 rpm in the present embodiment) and a minimum speed of rotation (1000 rpm) for the operating points

-   -   suction pressure ps=35 bar; compression pressure pd=130 bar     -   suction pressure ps=50 bar; compression pressure pd=100 bar and     -   suction pressure ps=20 bar; compression pressure pd=70 bar.

The regulation characteristic curves (not shown) of a speed of rotation located between 1000 rpm and 8000 rpm lie between the respective curves having the particular suction and compression pressures. As can be seen from FIG. 3, the regulation characteristic curves for the speed of rotation 8000 rpm for the three different operating points intersect the x axis (on which there is plotted the tilt angle of the tilt plate or the geometric stroke volume of the compressor) at about half the maximum tilt angle, which in the case of the present preferred embodiment is about 20°. For exactly one speed of rotation, the intercept with the x axis can be adjusted by structural means; the other operating points have minor differences therefrom, with regulation characteristic curves for the same speed of rotation and different operating points being displaced approximately parallel to one another.

FIG. 4 shows regulation characteristic curves for one operating point (suction pressure ps=35 bar; compression pressure pd=130 bar) for various speeds of rotation, namely 1000 rpm, 2000 rpm, 4000 rpm and 8000 rpm, the spring constant of the restoring spring being 30 N/mm. As can be seen from the Figure, for speeds less than or equal to 4000 rpm it is possible to reach an angle of maximum deflection of about 20° without any problem, whereas at 8000 rpm it is now possible only to reach an angle of maximum deflection of about 14°. Accordingly, up to speeds of rotation of 4000 rpm, possibly even 5000 rpm, it is possible to reach the maximum piston stroke, whereas for higher speeds of rotation the angle of greatest deflection of the tilt plate is reduced. By increasing the spring constant of the restoring spring or also, however, by increasing the mass inertia of the tilt plate in relation to the piston mass, it would also be possible to obtain a compressor wherein at a speed of rotation of 4000 rpm it would already be no longer possible to reach the maximum piston stroke.

The above-described tendency can be seen from FIG. 5, which shows the regulation characteristic curves for the same operating points as in FIG. 4 but for a spring constant of 60 N/mm. As can be clearly seen, the curves shown indicate a behaviour which favours down-regulation of the compressor. In the context of the present invention it was recognised that at the operating point in this case, comprising a suction pressure of ps=35 bar and a compression pressure pd=130 bar, an increase in the spring constant from 30 N/mm to 60 N/mm gives rise to a regulation characteristic curve which is somewhat more advantageous than in the case of a lower spring constant. On the other hand, as was recognised in accordance with the invention, for other operating points, for example a suction pressure ps=20 bar and a compression pressure. pd=70 bar (cf. FIG. 3), an increase in spring constant is disadvantageous, because the regulation characteristic curves would have a greater degree of slope. The intercept with the x axis would already happen at relatively low speeds of rotation. To summarise the above explanations it can be concluded that it has been recognised in accordance with the invention that, for the restoring spring, spring rates of less than 60 N/mm, especially spring rates of about 30 N/mm or possibly.even less, are advantageous.

As an alternative to variation of the spring rate or also possibly of course in addition thereto, the mass inertia of the tilt plate can be increased. FIG. 6 in turn shows the regulation characteristic curves for a suction pressure ps=35 bar and a compression pressure pd=130 bar for the various speeds of rotation at a spring constant of 30 N/mm and with increased mass inertia of the tilt plate. As can be seen from FIG. 6, the desired course is again obtained. The details of the geometry of the tilt plate having a regulation characteristic curve as shown in FIG. 6 can be seen from Table 2 hereinbefore.

Both of the tilt plate geometries given in Tables 1 and 2 are suitable for obtaining the desired regulation behaviour, as can be seen from the Figures explained hereinbefore.

Finally, FIG. 7 shows, analogously to FIG. 3, the known three operating points at speeds of rotation of both 1000 rpm and 8000 rpm. It is desirable for a compressor to have an arrangement of regulation characteristic curves approximately in accordance with FIGS. 3 and 7 with the stated mass inertias of the tilt plates in accordance with Table 1 and Table 2. The regulation characteristic curves from the diagram of FIG. 3 show the desired effect in comparatively less pronounced form compared to the regulation characteristic curves from FIG. 7, the effect in the regulation curves according to FIG. 7 in contrast being much more pronounced. However, both embodiments have in common the fact that the regulation characteristic curves for high speeds of rotation have an intercept with the x axis, which means that the angle of maximum deflection for high speeds of rotation is less than the angle of maximum deflection for low speeds of rotation. It is especially desirable for the regulation characteristic curves for speeds of rotation above about 4000 rpm to have such an intercept, that is to say to have a reduced angle of maximum deflection of the tilt plate. In the case of high speeds of rotation it is desirable for the stroke volume to be limited to about 50% or less, with high speeds of rotation being about 8000 rpm in the case of modern compressors as already mentioned.

It should be mentioned that the down-regulation of the compressor would also be achievable by means of a regulatory intervention, in which case energy losses also play a large part in addition to the inertia over time which the regulatory intervention involves. In the case of strong acceleration of the vehicle and operation of the compressor under full load, compressor operation initially demands a very high torque (at least until the control corrects this again and the stroke volume is reduced), which is of course extremely undesirable especially during overtaking manoeuvres of the vehicle. In addition, safety is increased in a compressor according to the invention, because malfunctions of the control valve, for example due to dirt accumulation, blockage of the valve seat, icing-up or the like, impair or impede regulatory intervention. A marked pressure increase, especially in the case of high-pressure coolants such as CO₂, would very easily become a safety risk in such a case. It is therefore desirable that, at a high speed of rotation, relatively high tilt angles of the tilt plate are not achievable in the first place.

It should furthermore be noted that FIGS. 3, 4, 5, 6 and 7, which relate to compressors in which the high-pressure coolant CO₂ is used, can of course also be used qualitatively for other coolants such as R134a, R152a etc. It would be necessary merely to modify the y axes of the diagrams with regard to the stated pressures.

It should furthermore be pointed out that the desired regulation characteristic curve (having intercepts of the appropriate curves with the x axis) is basically achieved by suitable selection of M_(SW) and M_(k, ges). The fine adjustment of the desired regulation characteristic curve can also be accomplished by other measures such as, for example, suitably selecting the spring constant of the restoring spring.

Finally, reference should be made to FIG. 8, which shows the interplay between the tilt angle of the tilt plate, or swash plate, and the pressure in the drive mechanism chamber, or the differential pressure between the drive mechanism chamber and the intake chamber, in dependence on various parameters. From FIG. 8 there can be seen the regulation characteristic curves of a compressor according to the invention using various additional arrangements which exert an adjusting force on the tilt plate. In the case of the preferred embodiments of a compressor according to the invention discussed herein, the arrangements referred to above as additional arrangements are arrangements which are associated with the compressor besides an arrangement for adjustment or variation of the drive mechanism chamber pressure and in addition to a restoring spring. At this juncture it should be noted that, besides geometric considerations relating to moment distributions, other arrangements which influence or limit the tilt angle of the tilt plate, that is to say arrangements which, for example, exert an adjusting force on the tilt plate, are of course also included in the basic concept of the present invention. In detail, in addition to the already mentioned restoring spring, these can be an actuator, an adjusting piston, an adjusting member which is dependent on centrifugal force, a further spring or also a throttling location of variable cross-section or of constant cross-section. A throttling location of variable cross-section especially in the form of a throttling location which is variable by means of an adjusting member, that is to say by means of an adjusting piston, is feasible. It should be emphasised that the scope of the invention includes both single measures from the afore-mentioned measures and also a plurality of measures or arrangements in combination. The corresponding regulation characteristic curves can (as already mentioned hereinbefore) be found in FIG. 8.

In detail it can be seen from FIG. 8 that, by increasing the spring constant of the restoring spring or as a result of optionally using a further adjusting force, that is to say for example by using an actuator, a steeper regulation characteristic curve can be obtained. It can furthermore be seen from FIG. 8 that throttling on the intake side, for example by means of a throttling location of variable cross-section but also of fixed cross-section results in a curve which is displaced towards lower pressures, parallel to the regulation characteristic curve without the additional adjusting force. An analogous regulation characteristic curve is obtained when the restoring spring is biased. As a result of the measures described hereinbefore it is especially possible for the desired regulation characteristic curve having intercepts of the curve in question with the x axis at the desired location to be produced by simple means, the wide range of possibilities allowing a wide regulation spectrum (depending on the field of application of the compressor). Depending on the requirement of the user, one or more measures can be combined and accordingly any desired regulation characteristic curve obtained.

Although the invention is described using embodiments having fixed combinations of features, it nevertheless also encompasses any further feasible advantageous combinations of those features, as are especially but not exhaustively mentioned in the subordinate claims. All features disclosed in the application documents are claimed as being important to the invention insofar as they are novel on their own or in combination compared with the prior art. 

1. Axial piston compressor, especially a compressor for the air-conditioning system of a motor vehicle, having a housing and, for drawing in and compressing a coolant, a compressor unit arranged in the housing and driven by means of a drive shaft, the compressor unit comprising pistons, which move axially back and forth in a cylinder block, and a tilt plate (swash plate, wobble plate or tilt ring) which drives the pistons and rotates together with the drive shaft, characterised in that the tilting behaviour of the tilt plate acts in a self-limiting manner such that at high speeds of rotation of the compressor, especially at very high speeds of rotation or at the maximum speed of rotation, the angle of maximum deflection of the tilt plate is less than the angle of maximum deflection α_(max) at low speeds of rotation of the compressor.
 2. Compressor according to claim 1, characterised in that the geometry and dimensioning of all parts moved in translation, such as axial pistons, piston rods or sliding blocks or the like, on the one hand, and all parts moved in rotation, such as the tilt plate, members for conjoint movement or the like, on the other hand, are such that, for predetermined tilt angles of the tilt plate, especially between a predetermined minimum tilt angle and a predetermined maximum tilt angle, the moment M_(k,ges) due to the masses moved in translation, especially that of the pistons, where appropriate including sliding blocks, piston rods or the like, is so selected as to be less than the moment M_(SW) due to the moment of deviation, that is to say than the moment due to the mass inertia of the tilt plate, that at high speeds of rotation of the compressor, especially at very high speeds of rotation or at a maximum speed of rotation, the angle of maximum deflection of the tilt plate is less than the angle of maximum deflection α_(max) at lower speeds of rotation of the compressor.
 3. Compressor according to claim 1, characterised by at least one arrangement which exerts an adjusting force on the tilt plate.
 4. Compressor according to claim 3, characterised in that the at least one arrangement for exerting the adjusting force is associated with the compressor in addition to an arrangement for adjusting the pressure in the drive mechanism chamber.
 5. Compressor according to claim 3, characterised in that the at least one arrangement for exerting the adjusting force comprises an actuator or an adjusting piston.
 6. Compressor according to claim 3, characterised in that the at least one arrangement for exerting the adjusting force comprises an adjusting member which is dependent on centrifugal force.
 7. Compressor according to claim 3, characterised in that the at least one arrangement for exerting the adjusting force comprises a spring, especially in addition to a restoring spring.
 8. Compressor according to claim 3, characterised in that the at least one arrangement for exerting the adjusting force comprises a throttling location.
 9. Compressor according to claim 8, characterised in that the throttling location is arranged on the intake side of the compressor.
 10. Compressor according to claim 8, characterised in that the cross-section of the throttling location is variable by means of an adjusting member, especially an adjusting piston.
 11. Compressor according to claim 10, characterised in that the pressure (P_(c)) prevailing in the drive mechanism chamber of the compressor and/or the pressure (P_(d)) prevailing on the output side serve(s) as the adjusting variable for the adjusting member, especially the adjusting piston.
 12. Compressor according to claim 10, characterised in that the adjusting member is regulated or controlled internally, especially by means of a solenoid or like apparatus.
 13. Compressor according to claim 8, characterised in that the throttling location has a constant cross-section.
 14. Compressor according to claim 1, characterised in that the centre of gravity of the tilt plate is located on the tilt axis thereof.
 15. Compressor according to claim 14, characterised in that the tilt axis of the tilt plate is located on the central axis of the drive shaft.
 16. Compressor according to claim 1, characterised in that the tilt plate is ring-shaped, that is to say in the form of a tilt ring.
 17. Compressor according to claim 1, characterised in that the angle of maximum deflection of the tilt plate in the region of low and medium speeds of rotation up to about half the maximum speed of rotation of the compressor, especially in a range from 0 to 4000 revolutions per minute, corresponds to an angle α_(max), whereas at the maximum speed of rotation of the compressor it corresponds approximately to an angle of α_(max)/2.
 18. Compressor according to claim 1, characterised in that, for speeds of rotation from about 4000 revolutions per minute to 5000 revolutions per minute upwards, the tilt plate is no longer capable of being pivoted up to an/the angle amax.
 19. Compressor according to claim 1, characterised in that the limitation of the pivot angle becomes greater for increasing speeds of rotation of the compressor.
 20. Compressor according to claim 2, characterised in that the ratio of the moment of deviation in the y direction to the total mass of all parts moved in translation such as, for example, pistons and, where appropriate, sliding blocks and the like, J_(y)/m_(ges), is at least about 1000 gmm²/g, especially more than 1500 gmm²/g.
 21. Compressor according to claim 2, characterised in that the ratio of the moment of deviation in the z direction to the total mass of all parts moved in translation such as, for example, pistons and, where appropriate, sliding blocks and the like, J_(y)/m_(ges), is at least about 2000 gmm²/g, especially more than 3000 gmm²/g.
 22. Compressor according to claim 4, characterised in that a or the restoring spring has a spring constant of less than about 60 N/mm, especially less than 30 N/mm.
 23. Compressor according to claim 1, characterised by an end-stop which is dependent on the speed of rotation and which limits the angle of maximum deflection of the tilt plate in dependence on the speed of rotation of the compressor. 